Fuel efficient hydraulic power steering

ABSTRACT

A hydraulic power steering system that maintains proportional pressure control between an input side of a gear and assist pressures internal to the steering gear by relying upon a control valve. A pressure differential is maintained between the input side of a gear and assist pressures internal to the steering gear throughout certain events. Those events include running the pump to fully charge the accumulator, idling the pump after fully charging the accumulator, discharging pressure from the fully charged accumulator at an onset of a steering event having a demand for steering load, recovering the discharged pressure of the accumulator by running the pump after turning on the pump at the onset of the steering event, and opening the relief valve in the valve manifold during the steering event.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The invention relates to a hydraulic power steering system thatcontinuously provides a source for hydraulic assist to a steering gearwhile allowing the pump to be idled when there is no steering demand toreduce energy consumption.

2. Description of Related Art

Conventional hydraulic power assist steering systems employ an opencenter valve at the steering gear and a pump, which supplies constantflow through an open loop hydraulic circuit. Fluid is pumped through thesystem at all times while the engine is running regardless of steeringload. By closing ports within the steering valve creating pressureinternal to the gear and in the hydraulic circuit upstream to the pump,boost or steering assist is created. Running a hydraulic pumpcontinuously in this type of system ensures good response to a steeringinput but is inefficient when there is no steering load. Restrictions inthe hydraulic system can create significant backpressure against thepump and generate a significant amount of heat over time that must bemanaged with oil coolers. As a result, prior studies have shown thatfuel consumption in the unloaded mode can dominate the fuel consumedunder load.

U.S. Pat. No. 5,921,342 (the '342 patent) provides a Power AssistedSteering Apparatus for Automotive Vehicle, but does not have aproportional pressure relief for the pump under load but instead relieson throttling down flow through the system using a variable displacementpump when steering demand is low. Since the pressure drop acrossrestrictions in the hydraulic system is proportional to flow rate, thebackpressure that the pump needs to work against is lower when flow isreduced through the system in low steering demand situations. Therefore,this system reduces pumping losses and saves energy over hydraulic powersteering systems that employ fixed displacement hydraulic pumps.

The patent application publication to Rogers et al. (2003/0127275 A1)provides for a High Efficiency Automotive Hydraulic Power SteeringSystem and does have an automotive hydraulic power steering system thatprevents wasted energy when no power assist is required but does notmaintain proportional pressure control between the input side of thegear and the assist pressures internal to the gear. Proportional controlis a key enabler to minimizing leakage through the steering valve andbeing able to control steering assist as a function of steering valveopening hydraulically without the need for electronic controls. Thesystem described does not provide a load sensing signal downstream ofthe steering valve. Without a load sensing signal, the low pressuresetting for the hysteresis pressure switch must be raised to ensureadequate response during worst case steering conditions which reducesthe effective range of the accumulator.

SUMMARY OF THE INVENTION

One aspect of the invention resides in the design of a power steeringhydraulic system that includes an engine driven power steering pumpequipped with a clutch, a steering gear with a closed center controlvalve, a piston type accumulator, a valve control manifold and twopressure switches. The valve manifold provides pressure to the inputside of the closed center steering valve that is proportional to theassist pressure internal to the gear so that the level of assist is bothpredictable and consistent as a function of the steering valve opening.A shuttle valve is used in conjunction with a pressure switch as thesteering load sensing device. The accumulator is used as a storagedevice to permit the pump to be idled when there is no steering load andalso to provide uninterrupted assist to the gear during pump clutchengagement at the onset of a steering event.

The inventor established the following design scheme in developing theinvention with the objective of maximizing energy efficiency ofhydraulic power steering systems—particularly those with higher steeringload requirements.

Utilize the existing power source to drive the steering pump. Adding aseparate power source (typically an electric motor) for power steeringhas clear advantages as it can be optimized for the application—ElectricPower Steering (EPS) and Electro-Hydraulic Power Steering (EHPAS) aretypical examples. However, the size and cost of electric motors neededfor higher steering load applications make these systems undesirable orimpractical at this point in time.

Turn off the pump when there is no steering load—As stated earlier,prior studies have shown that fuel consumption in the unloaded mode candominate the fuel consumed under load. Therefore, this is a key tominimizing fuel consumption.

Use a closed center steering gear valve to eliminate fluid flow throughthe system when there is no steering load. Conventional open centervalves require constant flow at all times. A closed center valve allowsfluid to be stored upstream in the system and delivered to the gear ondemand.

Add an accumulator to manage steering load/unload transitions. Thisenables initial steering assist (and/or brake assist if equipped withhydra boost) while the pump is being engaged at the onset of a steering(and/or braking) event. Prior studies with clutched pumps withoutaccumulators have shown unacceptable response. The closed centersteering gear valve provides flow on demand and in proportion tosteering wheel rate rather than continuous flow, which minimizesaccumulator size.

Use pressure internal to the gear and a switch as the steering loadsignal. Load sensing with a closed center steering valve can also beaccomplished by sensing pressure drop across an orifice in the supplyline to the gear or with a steering angle sensor on the gear inputshaft. However, both load sensing signals have limitations. The pressuredrop across an orifice can be difficult to measure when steering wheelrates (and consequently flow to the gear) are low but steering load andneed for assist are high. Steering angle alone does not always correlatewell to steering load and often requires other inputs, signal processingand calibration complications.

Use proportional hydraulic valve controls to control assist and pumpflow. Proportional control allows supply and demand to balancehydraulically without the need for complicated and costly electroniccontrol strategies.

BRIEF DESCRIPTION OF THE DRAWING

For a better understanding of the present invention, reference is madeto the following description and accompanying drawings, while the scopeof the invention is set forth in the appended claims.

FIG. 1 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, at the start of the accumulator chargecycle with no steering load.

FIG. 2 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting the end of the accumulatorcharge cycle with no steering load.

FIG. 3 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting the onset of a light steeringevent.

FIG. 4 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting the onset of a steering eventthat triggers pump clutch engagement.

FIG. 5 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting accumulator fully charged duringa steering event.

FIG. 6 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting when the relief valve in thevalve manifold begins to open during a steering event.

FIG. 7 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting when the relief valve in thepump begins to open during a max load steering event.

FIG. 8 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting release of the steering wheelafter a max load steering event.

FIG. 9 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting the pressure to the steeringgear valve returning to set point through leakage at the steering gearvalve.

FIG. 10 shows the hydraulic circuit diagram in accordance with anembodiment of the invention, reflecting accumulator pressure drop asflow is supplied to gear.

FIG. 11-13 show the electrical circuit diagram in accordance with thesame embodiment of the invention and operation during the hydraulicpower steering system events described in FIGS. 1-10.

FIG. 14-15 show the function of respective pressure switches asreferenced in the hydraulic circuit diagrams shown in FIGS. 1-10.

FIG. 16 shows a free body diagram of the relief valve of FIGS. 1-10.

FIGS. 17-18 show the relief valve of FIG. 16 in operation foraccumulator charging, respectively, at a beginning of the accumulatorcharging cycle and at an end of the accumulator charging cycle.

FIGS. 19-20 show the relief valve of FIG. 16 in operation for a steeringevent, respectively, closed and open.

FIG. 21 shows a response curve describing the function of control valvereferenced in the hydraulic circuit diagrams shown in FIGS. 1-10.

FIG. 22 shows a response curve describing the function of relief valvereferenced in the hydraulic circuit diagrams shown in FIGS. 1-10.

DETAILED DESCRIPTION OF THE INVENTION

FIGS. 1-10 share the following components in the hydraulic powersteering system: a pump/clutch assembly 10, an accumulator 12, a brakehydra boost 14 with brake master cylinder 16, a steering gear 19, asolenoid valve 20 that is normally open, a relief valve 22 that isnormally closed, a closed centered, (gear valve or) steering valve 18, ashuttle valve 24, a control valve 26 that is normally open, a checkvalve 28 downstream of the pump/clutch assembly 10, a pressure switch 32in the accumulator circuit, a pressure switch 34 in the referencepressure circuit downstream of shuttle valve 24 and reservoir 15.

The clutch of the pump/clutch assembly 10 engages the pump when there isa steering demand or accumulator needs to be charged. Shuttle valve 24monitors pressure on each side of the piston in steering gear 19 andprovides the higher of the two pressures to the reference pressurecircuit. Accumulator 12 is used to store power steering fluid underpressure. Flow from accumulator 12 allows control valve 26 to provideinitial pressure/flow to steering gear 19 through steering valve 18 toprevent a response lag because of time delay from engagement of thepump/clutch 10 until the pump runs at applicable capacity during asteering event.

Solenoid valve 20 (normally open) allows relief valve 22 to close morerapidly when the clutch disengages after an accumulator charge. Thesolenoid in solenoid valve 20 energizes and closes the valve when thepump clutch is engaged blocking flow to reservoir 15 and allowing pumppressure to be sensed by relief valve 22. The solenoid in solenoid valve20 de-energizes and opens the valve when the pump clutch is disengagedallowing flow through solenoid valve 20 to reservoir 15. This allowspressure to relief valve 22 to be vented to atmosphere driving the valveclosed immediately after the pump clutch is disengaged since reliefvalve 22 is normally closed.

Relief valve 22 is normally closed and begins to open when pump supplypressure exceeds the pressure required to overcome the spring settinginternal to relief valve 22 plus a secondary opposing pressure from theopposite end of relief valve 22 which is ported to the referencepressure circuit. A unique feature of relief valve 22 is that the areaof the internal spool valve that is exposed to reference pressure is 30%smaller than the area of the spool valve that is exposed to the pumpsupply pressure. This configuration reduces the contribution ofreference pressure on the spool and allows relief valve 22 to beginrelieving some of the flow to the pump back to reservoir 15 as referencepressure increases to better balance pump supply with demand asdetermined by reference pressure. This valve design keeps the pump fromgoing into relief prematurely in moderate to high steering loadconditions. The operation of relief valve 22 is shown graphically inFIG. 16.

Schematic diagrams of the on-demand hydraulic power steering system inaccordance with an embodiment of the invention are shown in FIGS. 1-10,each at different operating conditions. Color is used in the diagrams torepresent the pressure in the circuit at each operating condition—greenrepresenting the lowest pressure and red the highest. In ascending orderof pressure, the colors used in FIGS. 1-10 are green, tan/brown, yellow,orange, red.

FIG. 1 shows the hydraulic circuit diagram at the start of theaccumulator charging cycle when there is no steering load (or brake loadif the application includes hydra boost braking system). The pressure inaccumulator 12 has been dropping due to small but expected leakagethrough steering valve 18, which is currently closed and on center(i.e., no steering input by the driver).

In this example, pressure switch 32 is designed to close when pressurein the accumulator drops below the switch setting. In this example,switch 32 has just closed providing power to pump clutch 10 and solenoidvalve 20. Power to the normally open solenoid valve 20 closes the valve.As the pump clutch 10 engages, flow from the pump opens check valve 28and begins to fill accumulator 12 while pressure builds in theaccumulator circuit. Shuttle valve 24 is idle since there is no steeringinput from the driver and is supplying zero pressure to the referencepressure circuit. At this point, relief valve 22 is closed.

As pressure builds in the pump supply and accumulator circuits, reliefvalve 22 remains closed allowing all pump flow to charge accumulator 12.The purpose of control valve 26 is to maintain proportional pressurecontrol between the input side of the gear and reference pressure. Inthis example, control valve 26 is set to continuously maintain apressure to the input side of the gear (ex. 300 psi) proportional to thevalve spring setting. The operation of control valve 26 is showngraphically in FIG. 21.

FIG. 2 shows the hydraulic circuit diagram of the same embodimentreflecting the end of the accumulator charging cycle when there is nosteering load (or brake load if application includes hydra boost). Thepressure in accumulator 12 has been rising with the pump clutch 10engaged.

In this example, pressure in the accumulator circuit has reached themaximum accumulator charge pressure and triggered the pressure switch 32to open which turns off electrical power to pump clutch 10 and solenoidvalve 20. As pressure in the accumulator circuit approaches the triggerpoint of pressure switch 32, relief valve 22 also limits pressure in thecircuit during accumulator charging and begins to open when pump supplypressure reaches the maximum accumulator charging pressure (Thereference pressure contribution in this case is zero since there is nosteering load).

As pump clutch 10 disengages, flow from the pump stops, solenoid valve20 opens, check valve 28 closes to prevent back flow and pressureimmediately upstream of the check valve 28 is released to a ventedreservoir 15. Opening solenoid valve 20 allows pump supply referencepressure to relief valve 22 to drop and the valve to close immediatelyto prevent fluid and pressure loss from the accumulator circuit.Although pressure in the accumulator circuit is now at maximum chargingpressure, control valve 26 continues to maintain a set pressure (ex. 300psi) to the input side of the steering valve 18, which is still closedand on center. Keeping the pressure low to steering valve 18 minimizesleakage through the valve, maximizes the time between accumulatorcharges and subsequently maximizes fuel efficiency.

FIG. 3 shows a hydraulic circuit diagram of the same embodimentreflecting when a light steering event begins. When the driver turns thesteering wheel, the column shaft, the intermediate shaft between columnand steering gear 19 turn as well (not shown). If the torque applied tothe steering wheel is sufficient to overcome the torsion rate of a smallsteel bar (often referred to as a T-bar) hard mounted between the gearinput shaft and valve sleeve, the steering valve 18 will begin to open.In FIG. 3, the light steering event is represented, i.e., by showinghigher pressure on one side of steering gear 19 as steering valve 18begins to open.

The pressure internal to the left side of the steering gear 19 hasincreased (ex. 150 psi) and that the ball in shuttle valve 24 has beenforced to the right blocking flow to the other side of the gear. Shuttlevalve 24 then feeds the full pressure internal to the left side of thesteering gear to the reference pressure circuit. Since the piston insidegear 19 is tied directly to the vehicle road wheels through tie rods(not shown), the piston begins to move right in this case to turn thewheels.

As the piston moves to the right inside the gear housing, control valve26 must provide flow through steering valve 18 to fill the increasingvolume to the left of the gear piston and maintain proportional pressurecontrol between the input side of the steering valve 18 and thereference pressure. The control valve 26 has increased supply pressureto the steering valve 18 (ex. 450 psi) to a level equal to the presetvalve spring setting (ex. 300 psi) plus the reference pressure (ex. 150psi). The accumulator 12 supplies flow passing through control valve 26to the steering gear 19 and consequently the accumulator pressure drops(ex. 900 psi). The pressure in the accumulator circuit is betweenminimum and maximum accumulator charging levels so pressure switch 32remains open so there is no power to the pump clutch. The referencepressure (ex. 150 psi) is below the trigger point for pressure switch 34(ex. 300 psi), so the switch 34 remains open as well so the clutch tothe pump is disengaged.

FIG. 4 shows the hydraulic circuit diagram of the same embodimentreflecting when a steering event triggers pump clutch engagement. As thedriver continues a steering event (from FIG. 2) and applies more torqueto the steering wheel to steer the vehicle, steering valve 18 opensfurther allowing more flow and increases pressure to the left side ofthe steering gear 19. Control valve 26 continues to maintainproportional pressure between the input side to steering valve 18 andreference pressure. Reference pressure is now above the trigger pointfor pressure switch 34 (ex. 300 psi), so the switch has just closedsending power to the pump clutch. Check valve 28 now opens as pump speedincreases.

The control valve 26 continues to allow flow to the input side ofsteering valve 18 as the piston in the steering gear 19 moves right andhas increased supply pressure (ex. 600 psi) proportionately abovereference pressure (ex. 300 psi). Accumulator pressure has dropped (ex.700 psi) and must have sufficient capacity to supply the input side ofsteering valve 18 while the pump clutch is engaging. Depending oncomponent selection and system layout of the hydraulic system, pressureswitch 32 and 34 settings may need to be adjusted to ensure adequatepressure and flow can be provided by control valve 26 to the input sideof the steering valve 18 while the pump clutch is engaging during rapidsteering inputs.

FIG. 5 shows the hydraulic circuit diagram of the same embodimentreflecting recovery of accumulator pressure during a steering event withpump clutch engaged. Since pumps are sized to deliver enough flow atengine idle to meet worst case steering requirements, output flow fromthe pump/clutch assembly 10 will exceed flow required at gear 19 undermost conditions. The excess flow will then begin building pressure inthe accumulator circuit.

In this case, pressure in accumulator 12 has fully recovered to the maxaccumulator pressure setting. Control valve 26 continues to increasepressure (ex. 1200 psi) to the input side of steering valve 18 inproportion to reference pressure (ex. 900 psi). The piston in gear 19continues to move to the right as the driver continues to turn thesteering wheel and steering load and assist pressures are increasing.Pressure switch 34 and solenoid valve 20 are closed. Relief valve 22remains closed because the pump supply pressure (ex. 1300 psi) is stillbelow the proportional cracking pressure of relief valve 22 (ex. 1570psi=1300 psi spring+0.3×900 psi reference pressure).

FIG. 6 shows the hydraulic circuit diagram of the same embodimentreflecting the relief valve in the manifold beginning to open during asteering event. As the steering event continues, control valve 26continues to increase pressure (ex. 1470 psi) to the input side ofsteering valve 18 in proportion to reference pressure (ex. 1170 psi).Also, as the pump continues to run, check valve 28 is open, and thepressure continues to rise in the accumulator circuit. Eventually pumpsupply pressure reaches the cracking pressure of relief valve 22.

At this point, pump supply pressure (ex. 1600 psi) is nearly equal tothe cracking pressure of relief valve 22 (ex. 1651 psi=1300 psispring+0.3×1170 psi reference pressure). At this point, relief valve 22begins to open allowing excess pump flow to return to reservoir 15,which is vented to atmosphere.

FIG. 7 shows the hydraulic circuit diagram of the same embodimentreflecting the relief valve in the pump beginning to open during amaximum load steering event. Typically, maximum steering loads occurduring a parking type maneuver when the steering wheel is turned all theway to the end of travel (steering stops).

In FIG. 6, relief valve 22 was beginning to open. However, as flow isdirected to reservoir 15, control valve 26 continues to increasepressure (ex. 1850 psi) to the input side of steering valve 18 inproportion to reference pressure (ex. 1850 psi) as steering loadsapproach maximum levels. All pumps have internal relief valves toprotect the hydraulic system. In the example shown in FIG. 7, the reliefvalve internal to the pump is set to open at 1850 psi and it begins toopen redirecting pump flow within the pump itself. As the flow drops,check valve 28 closes. At this point, relief valve 22 is also beginningto open, because the proportional cracking pressure of relief valve 22(ex. 1855 psi=1300 psi spring+0.3×1850 psi reference pressure). Reliefvalve 22 provides a redundant fail safe to protect the system should thepump relief fail.

FIG. 8 shows the hydraulic circuit diagram of the same embodimentreflecting release of steering wheel (steering load to zero) after amaximum load steering event. When the steering wheel is released,steering valve 18 returns to the on-center position. On center, steeringvalve 18 is designed to allow flow between left and right side gearports in gear 19 and reservoir 15 through a return port in the valve. Assuch, when the steering wheel is released pressure on each side of thepiston in the steering gear 19 along with reference pressure throughshuttle valve 24 are relieved to atmospheric pressure. This drivescontrol valve 26 closed since it is designed to allow flow only whensupply pressure to input side of steering valve 18 is less thanreference pressure while the pressure differential at this point is muchhigher (ex. 1750 psi). Relief valve 22 closes since the pressure in thepump supply is zero and well below the cracking pressure of relief valve22 (ex. 1300 psi=1300 psi spring+0.3×0 psi reference pressure).

FIG. 9 shows the hydraulic circuit diagram of the same embodimentshowing pressure to steering valve 18 returning to the set point incontrol valve 26 (ex. 300 psi). This is accomplished by leakage throughsteering valve 18, which is on center.

FIG. 10 shows the hydraulic circuit diagram of the same embodimentreflecting a steering event in the opposite direction. In this example,the steering event is light similar to the event described in FIG. 2.Reference pressure (ex. 150 psi) is still below the trigger point forpressure switch 34 so the electrical circuit to the pump/clutch 10remains open. Note, that control valve 26 has opened again to provideflow to gear 19 and maintain proportional pressure control to steeringvalve 18.

FIGS. 11-13 show three different modes for the electrical circuit forthe same embodiment, which includes a battery that provides power toelectrical components, pump/clutch assembly 10, solenoid valve 20,pressure switch 32, and pressure switch 34.

FIG. 11 reflects the standby mode where both pressure switch 32 and 34are open. Since there is no path to ground for pump/clutch 10 andsolenoid valve 20 there is no power to either component. The pump/clutchassembly 10 in FIG. 11 is disengaged. FIG. 12 reflects the accumulatorcharge mode where pressure switch 32 is closed and pressure switch 34 isopen. Since the pressure switches are wired in parallel, there is a pathto ground providing power to pump/clutch 10 and solenoid valve 20. Thepump/clutch assembly 10 in FIG. 12 is engaged. FIG. 13 reflects asteering event. The effect is the same as for FIG. 12, except that thepressure switch 32 is now open while switch 34 is closed. Thepump/clutch assembly 10 in FIG. 13 is engaged.

FIG. 14-15 show the function of the pressure switch 32 and the pressureswitch 34 for the same embodiment. FIG. 14 plots the function of switchposition versus accumulator pressure for the pressure switch 32 whileFIG. 15 plots the function of switch position versus reference pressurefor the pressure switch 34. The hydraulic pressures at which theswitches open and close vary depending on the application. The solidline arrows represent the response to increasing pressure, while thedashed line arrows represent the response to decreasing pressure.

FIGS. 16-20 show the relief valve 22 in various modes of operation. FIG.16 shows the relief valve 22 in a free body diagram. Directional arrowsare depicted to show the force direction for the F(pump), F(reference)and F(spring).

The force balance equations for the relief valve 22 are:

F(pump)=F(reference)+F(spring)

Since Pressure=Force/Area,

P(pump)*A1=P(reference)*A2±F(spring), with A1 & A2 defining the crosssectional areas at each end of the spool valve according to the relationA2=A1*0.3 so that:

P(pump)*A1=P(reference“A1*0.3+F(spring).

If the spring force is set so that F(spring)=P(accumulator fillsetting)*A1, then P(pump)*A1=P(reference)*A1*0.3+P(accumulator fillsetting)*A1

-   -   By dividing both sides by A1,

P(pump)=P(reference)*0.3+P(accumulator fill setting)

FIGS. 17 and 18 depict operation of the relief valve 22 during thecharging cycle of the accumulator 12 of FIGS. 1-10. FIG. 17 representsthe force balance on the relief spool during the beginning of theaccumulator charging, while FIG. 18 represents the force balance on therelief spool during the end of the accumulator charging.

-   -   The force balance equations for the relief valve 22:

Since the pump & accumulator are in the same circuit,P(pump)=P(accumulator).

-   -   If there is no steering event during the accumulator charging,        P(reference)=0.

Given the force balance equation:

P(pump)=P(reference)*0.3+P(accumulator fill setting), forces on thespool are balanced when P(pump)=P(accumulator fill setting). The valvewill remain closed until P(pump)>P(accumulator fill setting).

FIGS. 19 and 20 depict operation of the relief valve 22 during asteering event, with FIG. 19 showing the relief valve 22 in a closedposition and FIG. 20 showing the relief valve 22 in an open position.When the pump clutch is engaged, the pump delivers constant flow to thesystem. Flow through a steering gear (with a closed center steeringvalve 18 of FIGS. 1-10) varies based on the steering wheel rate and gearsize. The function of the relief valve is to dump excess flow from thepump to a reservoir. The relief valve uses the pressure differentialbetween the pump {P(pump)} and gear {P(reference)} to control flow.

For example, the force balance on the relief spool during a steeringevent:

If P(accumulator fill setting)=1300 psi and P(reference) at a point intime=1170 psi, then the force balance on the relief spool is achievedwhen P(pump)=1651 psi. That is:

P(pump)=P(reference)*0.3+P(accumulator fill setting)=1170 psi*0.3+1300psi=1651 psi

-   -   Consequently: if P(pump) rises above 1651 psi, the relief valve        22 opens, but if the P(pump) falls below 1651 psi, the relief        valve 22 closes.

FIG. 21 shows a graphical illustration of the function of the controlvalve for the same embodiment. The graph includes two lines: A referenceline (solid line) showing when the supply pressure to the gear equalsreference pressure and a dashed line that represents the pressureprovided by the control valve to the gear as a function of referencepressure. The dashed line describing the control valve function alsoshows the condition when the pump goes into relief. Increasing the valvepressure setting above the gear reference pressure increases leakagethrough the steering valve, decreases the time between successiveaccumulator cycles, may improve system response and/or reduce catch, andforces the pump into relief at lower steering load conditions.

FIG. 22 shows a graphical illustration of the function of the reliefvalve for the same embodiment. The graph includes two lines: A referenceline (solid line) showing when accumulator pressure equals referencepressure and a dashed line that represents the pressure provided by therelief valve on the accumulator circuit as a function of referencepressure. Decreasing the area of the valve spool exposed to thereference pressure relative to the spool area exposed to the accumulatorpressure flattens the slope of the line and prevents the pump from goinginto relief too early, allows higher accumulator storage pressuresettings than would otherwise be the case, increases the time betweenaccumulator cycles when there is no steering demand, reduces energyconsumption.

While the foregoing description and drawings represent the preferredembodiments of the present invention, various changes and modificationsmay be made without departing from the scope of the present invention.

1. A method of hydraulic power steering, comprising; maintainingproportional pressure control between an input side of a steering valvein a valve manifold and a steering gear with a control valve in a valvemanifold by maintaining a pressure differential between the input sideof the steering valve and the steering gear during events that includerunning a pump to fully charge an accumulator, idling the pump afterfully charging the accumulator, discharging pressure from the fullycharged accumulator at an onset of a steering event having a demand forsteering load, recovering the discharged pressure of the accumulator byrunning the pump after turning on the pump at the onset of the steeringevent, and opening a relief valve in the valve manifold during thesteering event.
 2. The method of claim 1, wherein the running of thepump occurs as pressure within the accumulator falls below a lowpressure limit, idling of the pump occurs as the pressure within theaccumulator exceeds a high pressure limit, the low pressure limit beinglower than the high pressure limit.
 3. The method of claim 1, whereinthe discharging of the accumulator at the onset of the steering eventprevents a lag in response to satisfying the demand for steering load,the lag otherwise arising because of a time delay before the pump, afterbeing turned on at the onset of the steering event, reaches capacity tosatisfy the demand for steering load.
 4. A method of claim 1, furthercomprising opening the steering valve in correspondence with turning ofthe steering gear and closing upon cessation of the turning of thesteering gear.
 5. A method of claim 1, further comprising monitoring twopressures one on each side of a piston in the steering gear, themonitoring taking place with a shuttle valve that provides a higher ofthe two pressures to a reference pressure circuit.
 6. A method of claim1, further comprising sensing pressure with a pressure switch, thepressure switch closing upon sensing a drop in pressure in the referencepressure circuit to a low pressure limit, the closing by the pressureswitch allowing power to reach a clutch that engages the pump to triggerpumping, the pressure switch opening upon sensing pressure in thereference pressure control reaching a high pressure limit, the openingof the pressure switch stopping the power to the clutch, which resultsin the clutch disengaging the pump to stop the pumping.
 7. A method ofclaim 4, further comprising a relief valve that begins to open as pumpsupply pressure exceeds pressure required to overcome a spring settinginternal to the relief valve plus a secondary opposing pressure from anopposite end of the relief valve that is ported to the referencepressure circuit.
 8. A method of claim 7, further comprising exposing anarea of an internal spool valve to reference pressure that is 30%smaller than an area of the spool valve that is exposed to the pumpsupply pressure, which reduces a contribution of the reference pressureon a spool and allows the relief valve to begin relieving some flow tothe pump back to a reservoir as the reference pressure increases, whichfurthers balancing pump supply with the demand as determined by thereference pressure and keeps the pump from going into relief prematurelyduring conditions of the steering load that are moderate to high, whichis higher than at the onset of the steering event that is light.
 9. Amethod of claim 1, further comprising idling the pump during times whenthere is no demand for the steering load from the steering wheel and nodemand for braking load from a hydra boost braking system downstream ofthe accumulator.
 10. A method of claim 1, wherein the pump has aninternal relief valve that opens during a maximum load steering eventthat arises during a parking maneuver when a steering wheel is fullyturned to an end of travel so that pump flow redirects within the pumpitself to cause a reduction in flow leaving the pump, closing a checkvalve in response to the reduction in flow to prevent reverse flow. 11.A method of claim 5, wherein the steering valve returns to a centerposition upon release of a steering wheel that causes the steering loadto be zero, the center position allowing flow between opposite gearports in the steering gear and a reservoir through a return port in thesteering valve so that pressure on each side of the piston in thesteering gear along with reference pressure through the shuttle valveare relieved to atmospheric pressure.
 12. A method of claim 11, furthercomprising leakage through the steering valve that is in the centerposition, returning pressure to the steering valve to set point in thecontrol valve because of the leakage, the set point being the amount ofthe pressure differential.
 13. An apparatus suited for hydraulic powersteering, comprising; a pump, an accumulator upstream of the pump, asteering valve in a valve manifold, a relief valve in the valvemanifold, the valve manifold being upstream of the pump, a steering gearoperative to turn the steering valve, and a control valve arranged inthe valve manifold to maintain proportional pressure control between aninput side of the steering valve in the valve manifold and the steeringgear by maintaining a pressure differential between an input side of thesteering valve and the steering gear during events that include runningthe pump to fully charge the accumulator, idling the pump after fullycharging the accumulator, discharging pressure from the fully chargedaccumulator at an onset of a steering event having a demand for steeringload, recovering the discharged pressure of the accumulator by runningthe pump after turning on the pump at the onset of the steering event,and opening the relief valve in the valve manifold during the steeringevent.
 14. The apparatus of claim 13, further comprising a pressureswitch that responds to pressure within the accumulator falling below alow pressure limit to trigger running of the pump, the pressure switchresponds to pressure within the accumulator exceeding a high pressurelimit to trigger idling of the pump, the low pressure limit being lowerthan the high pressure limit.
 15. The apparatus of claim 13, wherein theaccumulator is configured to discharge at the onset of the steeringevent to prevent a lag in response to satisfying the demand for steeringload, the lag otherwise arising because of a time delay before the pump,after being turned on at the onset of the steering event, reachescapacity to satisfy the demand for steering load.
 16. An apparatus ofclaim 13, wherein the steering valve is arranged to open incorrespondence with turning of the steering gear and to close uponcessation of the turning.
 17. An apparatus of claim 13, furthercomprising a clutch operative to engage the pump to turn on of the pumpto run the pump and to disengage from the pump to turn off the pump tostop the pump from continuing to run.
 18. An apparatus of claim 13,further comprising a shuttle valve configured to monitor two pressuresone on each side of a piston in the steering gear and to provide ahigher of the two pressures to a reference pressure circuit.
 19. Anapparatus of claim 18, further comprising a pressure switch arranged toclose upon sensing a drop in pressure in the reference pressure circuitto a low pressure limit and thereby allow power to reach a clutch thatengages the pump to trigger pumping, the pressure switch beingconfigured to open upon sensing pressure in the reference pressurecontrol reaching a high pressure limit to thereby stop the power to theclutch, which results in the clutch disengaging the pump.
 20. Anapparatus of claim 18, further comprising a relief valve that begins toopen as pump supply pressure exceeds pressure required to overcome aspring setting internal to the relief valve plus a secondary opposingpressure from an opposite end of the relief valve that is ported to thereference pressure circuit.
 21. An apparatus of claim 20, furthercomprising an area of an internal spool valve being exposed to referencepressure that is 30% smaller than an area of the spool valve that isexposed to the pump supply pressure, which reduces a contribution of thereference pressure on a spool and allows the relief valve to beginrelieving some flow to the pump back to a reservoir as the referencepressure increases, which furthers balancing pump supply with the demandas determined by the reference pressure and keeps the pump from goinginto relief prematurely during conditions of the steering load that aremoderate to high, which is higher than that at the onset of the steeringevent that is light.
 22. An apparatus of claim 13, further comprising ahydra boost braking systems downstream of the accumulator, the pumpbeing directed to idle while there is no demand for the steering loadfrom a steering wheel and no demand for a braking load by the hydraboost braking system.
 23. An apparatus of claim 13, wherein the pump hasan internal relief valve that opens during a maximum load steering eventthat arises during a parking maneuver when a steering wheel is fullyturned to an end of travel so that pump flow redirects within the pumpitself to cause a reduction of flow leaving the pump; a check valveconfigured to close in response to the reduction of flow to preventreverse flow.
 24. An apparatus of claim 18, wherein the steering valvereturns to a center position upon release of a steering wheel thatcauses the steering load to be zero, the center position allowing flowbetween opposite gear ports in the steering gear and a reservoir througha return port in the steering valve so that pressure on each side of thepiston in the steering gear along with reference pressure through theshuttle valve are relieved to atmospheric pressure.
 25. An apparatus ofclaim 24, further comprising wherein the steering valve is configured toallow leakage through the steering valve as the steering valve is in thecenter position, the control valve being configured to return pressureto set point because of the leakage, the set point being the amount ofthe pressure differential.